This invention relates to steam turbines and, more particularly, to a method and apparatus for admitting steam into a steam turbine to improve turbine efficiency and at the same time reduce equipment costs relating thereto.
In general, steam flow is directed into large turbines through multiple arcuate nozzle chambers disposed circumferentially in both upper and lower turbine casings. Steam admission into the nozzle chambers is regulated by valves grouped in steam chests with the valves opening to admit steam from the steam chests into the nozzle chambers though "spaghetti" piping, and closing to obstruct the flow thereto. Variations in turbine design include full-arc admission units in which every first stage nozzle is active at all load conditions, and partial-arc units in which the number of active first stage nozzles is varied in response to load changes.
More efficient control of turbine output under varying load conditions has been realized by the partial-arc admission designs wherein the inlet nozzles are separated into discrete groups and are contained in individual chambers. Relatively high efficiency is attainable by sequentially admitting steam through individual nozzle chambers with a minimum of throttling, rather than by throttling the entire arc of admission. Typically, steam control valves modulate together when multiple valves are used to regulate steam flow into a single nozzle chamber.
According to well-established principles of thermodynamics, it is theoretically demonstrable that maximum turbine efficiency for changing load conditions can be attained from the use of an infinite number of valve points. A valve point is defined as a state of steam admission in which one or more of the valves are either in the completely open or completely closed position. Maximum efficiency would therefore require an infinite number of valves. As turbine load increases or decreases, valves would open or close to provide or subtract infinitesimal increments of steam flow to the nozzle chambers. Aside from the practical impossibility of providing an infinite number of valves with corresponding pipes, inlets, etc., a large number of valves is not economically feasible. Each additional turbine shell penetration with its inlet snout, each conduit and each valve, substantially increase the equipment cost. In practice, control valves for a given turbine element usually number from four to eight.
Efforts to achieve the optimum compromise between improved turbine efficiency and increasing capital cost for increasing numbers of valves typically focus on two aspects which are interrelated. One of these involves reducing the number of shell penetrations. The other focus is to maintain the optimum number of valve points utilizing various valve activation sequences.
With respect to the first of the foregoing considerations, a comparison of eight-valve designs, having eight inlets to the nozzle chambers, with a four-valve design having four valves and four inlet snouts, reveals that the shell and nozzle chamber diameters are larger on the eight-valve designs. This results in thicker pressure vessel walls, heavier bolting and more weight which, in turn, result in increased costs of the turbine shell, the steam chests and the inlet piping.
Some eight-valve designs typically have six nozzle chambers in two sizes, a large central chamber and two smaller outboard chambers in each of the upper and lower casings. These are supplied by two steam chests, one to the left and one to the right. The larger nozzle chambers have two inlets connected to separate lines from each steam chest. As mentioned hereinabove, the valves supplying such chambers with more than one inlet will typically modulate in unison. Other designs have utilized a Y fitting to connect the two lines from the separate steam chests to a single large inlet snout for each central chamber. On later designs, a T fitting has been used to join the separate lines to the large cover chamber. These fittings reduce the number of shell penetrations and make possible the same shell and nozzle chamber diameters as the four-valve turbine designs, but, these fittings are expensive. It would thus be desirable to eliminate such features.
Another equipment consideration is that in existing installations, the inlet pipes to the nozzle chambers in the upper casing pass downward from the steam chest, turn 180.degree., and then lead upward. These turns in the piping are much more expensive than straight piping, and the elbows increase pressure losses. A prior solution involves inverting one of the steam chests and supplying all of the nozzle chambers in the upper casing from the inverted steam chest. But, this solution would create unbalanced piping forces, producing a turning moment on the turbine shell.
The second consideration has resulted in sequencing of the activation of nozzle chambers in order to maximize the number of valve points, while at the same time, compensating for energy loss and stresses on turbine blades when passing inactive nozzle chambers. When rotating blades leave an active arc of admission, the steam flowing from the active nozzle chambers must work on the stagnant steam to set it in motion again, resulting in decreased efficiency or displacement loss. If the sequence of activating or de-activating nozzle chambers is such that there is never more than one inactive nozzle area in the inner circumference of the turbine, i.e., two or more adjacent inactive chambers, it is known as single shock operation. Double shock operation involves two interruptions of the active arcs of steam admission, and doubling the displacement loss, theoretically, with a commensurate degradation in the heat rate. Hence, there has been a preference for single shock operation heretobefore.
A prior invention for an improved method of activating individual nozzle chambers to maximize turbine efficiency is disclosed in U.S. Pat. No. 4,325,670 to Silvestri, assigned to Westinghouse Electric Corporation. That method involves a sequence of activating and deactivating six nozzle chambers of two different sizes to increase the number of valve points, thereby reducing heat rate, i.e., increasing efficiency. Half of the chambers are initially activated to produce a 50% arc of admission. It has been demonstrated empirically that arcs of admission below 50% result in poorer thermal performance and higher thermal stress. The remaining chambers are then sequentially activated and deactivated to provide, in combination, valve points at 62.5%, 75%, 87.5% and 100% admission.
The following table presents the effect of double versus single shock in turbine heat rate between partial admission arcs of 75% and 87.5%:
______________________________________ Throttle Flow Heat Rates, Btu/Kwh (KJ/Kwh) LB/hour (Kg/hour) Single Shock Double Shock ______________________________________ 3,330,000 (1510478) 7919 (8354.9619) 7920 (8356.0169) 3,260,000 (1478726.2) 7920 (8356.0169) 7923 (8359.1821) 3,200,000 (1451510.4) 7919 (8354.9619) 7922 (8358.127) 3,130,000 (1419758.6) 7913 (8348.6315) 7917 (8352.8518) ______________________________________
In fact, because the jet of steam from an active nozzle chamber spreads out in the clearance space between the nozzle and the rotating blade at the boundary between active and inactive arcs of admission, and because side leakage flow moves into the inactive zone, displacement loss accompanying double shock operation would be less than twice that of single shock operation, where the inactive arcs are kept relatively small. It is desirable to eliminate expensive design features if this can be done without substantially diminishing thermal performance.
Furthermore, certain analyses have established that the stresses and forces resulting from operation at 75% admission with double shock are no higher and are probably lower than stresses at 50% admission and single shock operation on turbines with 2400 psig throttle pressure and side entry control stage blade roots.